The present invention relates to an adsorption thermal storage device in a heat pump for augmenting heat rejection, operating over the 300.degree. to 425.degree. K. temperature range, and, more particularly, to a storage device and method incorporated into a spacecraft thermal-management/thermal-transport system for augmenting spacecraft heat rejection so as to be capable of storing more thermal energy than conventional phase change materials and also be lighter than pumped loop or vapor compression heat pump systems without thermal storage. That is, the present invention involves a heat-pump-adsorption storage device (HPASD) as part of a heat rejection system which provides significant thermal storage in a vapor compression heat pump.
Cyclic thermal loads require that a heating or cooling system be designed for the maximum load, or that a thermal storage device be installed to average the load if the temperature requirements are to be met. Many thermal management systems (heating, air conditioning, etc.) have very high peak-to-average thermal energy rejection duty cycles and are therefore prime candidates for thermal storage systems. However, these storage systems must be reliable and accommodate significant energy storage per unit mass of the storage system. Spacecraft applications have restrictions which include total system weight or volume, complexity, and reliability. It is often not desirable to increase the size of the primary system but this is usually the only effective technique, since thermal storage devices known today suffer from long-term performance problems. For example, phase change materials (PCM) exhibit incongruent melting; metal hydrides are heavy and they degrade due to fragmentation after repeated cycling. Other systems that depend on sensible heat storage are too large and too heavy.
Terrestrial applications use ice storage to handle large peak thermal loads found in churches and auditoriums; however, this method is slow and heavy. Spacecraft applications, which have cyclic thermal loads that must be rejected to space through a radiator system, present a major design problem. The spacecraft systems are very weight- and radiator-area-sensitive and at the same time have thermal loads that can range by a factor of 10 from base to peak. Current designs use PCMs that have the phase transition at the desired operating temperature. These PCMs have four major problems: (1) they are poor thermal conductors, and metal strips or ribbons are sometimes added to improve the heat transfer rate; (2) they tend to supercool and make temperature control very difficult; (3) they require a much larger radiator area to provide the needed cooling for recycling; and (4) salt systems melt incongruently and cause phase separation on repeated cycling.
At the 300.degree. K. (a temperature that is appropriate for spacecraft thermal management applications), several solid-liquid PCMs exist. However, these materials are slow to respond, have low thermal conductivity in the solid phase (so heat removal from these materials is a problem), and/or they have poor thermal stability. The typical approach of using metallic fibers or fins within the material to increase the solid phase thermal conductivity has met with limited success and has resulted in significant specific mass increases. Problems related to thermal stability and long-term life have yet to be solved for many compounds (such as Glauber's salt). At higher power system storage temperatures, molten-salt-type phase change materials exist. These molten salts do not, however, melt in the 300.degree. K. electronic cooler range. Gallium is a liquid metal with excellent thermal conductivity and stability in the 300.degree. K. range but Gallium has a limited storage capability of only 80 kJ/kg.
The use of heat pumps alone have shown some improvement, since they can raise the heat-rejection temperature which makes the radiator more efficient, but the impact is small since the heat pump compressor must be sized for the maximum load. Several thermally driven heat pump configurations, including hermetically sealed Rankine-powered vapor-compression systems, hermetically sealed free-piston Stirling-powered vapor-compression systems, absorption systems, and chemical/mechanical heat pumps, have been considered for use in spacecraft heat rejection systems. The vapor compression heat pump system has, however, performance and weight advantages over the Stirling, Brayton, or chemical (adsorption or absorption) heat pump systems. A chemical/mechanical heat pump system incorporates advantages of both the chemical and the mechanical vapor compression systems and has the added advantage of no moving parts.
Typical radiation-hardened radiators currently weigh approximately 20 kg/m.sup.2. The use of heat pumps requires additional electrical power which must be added to the heat pump system weight in addition to the weight of the radiator. Studies of terrestrial heat pumps, which were not designed to be lightweight, have shown that electrically driven terrestrial residential and commercial split-systems range in specific mass from 11 to 18 kg/kW-cooled. Actual spacecraft heat pumps have to be considerably lighter. For instance, a sliding vane rotary compressor Environmental Control Unit (ECU) for the LANTIRN electro-optical pod system has a specific mass of 3.8 kg/kW-cooled. Performance predictions have been developed for a scroll compressor ECU unit with specific mass estimates of 5.85 kg/kW-cooled. Representative calculations will demonstrate the area and weight savings achieved with the use of a heat pump. Radiative heat transfer in space can be accurately modeled, by using an effective space temperature of 227.degree. K. Also, by assuming a typical radiator emissivity of 0.8, the heat rejection from a spacecraft radiator can be calculated from the Stefan-Boltzmann equation, EQU q=.epsilon..sigma.(T.sub.rad.sup.4 -T.sub.s.sup.4) [1]
where:
q is the heat flux per unit area, W/m.sup.2 PA1 .sigma. is the Stefan-Boltzmann constant, W/m.sup.2 K PA1 .epsilon. is the emissivity of the radiator surface (0.8) PA1 T.sub.rad is the radiator temperature, K PA1 T.sub.s is the effective space temperature, 227.degree. K.
Values of the heat flux for several radiator temperatures are shown in Table I.
TABLE I ______________________________________ Radiator Heat Flux for Several Radiator Surface Temperatures Surface Temperature Radiator Heat Flux ______________________________________ 300.degree. K. 247 W/m.sup.2 320.degree. K. 560 W/m.sup.2 360.degree. K. 777 W/m.sup.2 375.degree. K. 1,358 W/m.sup.2 425.degree. K. 45,240 W/m.sup.2 ______________________________________
Assuming a 10 kW cooling-requirement base-line system that does not use a heat pump and therefore uses a 300.degree. K. radiator, the area required is 40.5 m.sup.2. If an electrically-driven heat pump with methanol as the working fluid (COP.sub.c of 1.85, compressor efficiency of 0.6, and condenser outlet temperature of 375.degree. K.) is used, then for a cooling requirement of 10 kW, the condenser energy rejection is 15.4 kW. It is useful to compare the savings in radiator area and total system weight that would result from the use of this heat pump system. The heat pump requires 5.4 kW of electrical energy. Assuming that the waste heat from the electrical power system is rejected from a high-temperature (1000.degree. K.) radiator and assuming an 8% electrical power conversion efficiency, 62.1 kW of additional waste heat must be rejected at 1000.degree. K. to supply this electrical energy. This translates into an additional 1000.degree. K. radiator area requirement of 1.4 m.sup.2.
The heat pump system will require 16.1 m.sup.2 of radiator area to reject the 15.4 kW of energy at 375.degree. K. Thus the heat pump results in a 37% reduction in the weight and a 60% reduction in the radiator area.
FIG. 1 displays a basic vapor compression heat pump system. The basic operation starts at the evaporator where a thermal load is balanced by evaporative cooling of the refrigerant to maintain a constant evaporator temperature. The vapor is compressed and heated before discharge to the condenser. This increase in temperature translates to more efficient radiator operation, which can more than offset the increased mass of the heat pump and associated electrical equipment. For example, a heat pump with methanol as the refrigerant and a condenser outlet temperature of 375.degree. K. can provide a 41% savings in mass and a 51% savings in radiator area over a pumped fluid loop.
When increased cooling is required, the thermostatic expansion valve (TXV) on the vapor compression heat pump system senses an increased evaporator exit superheated. This is because the refrigerant flow rate is insufficient for the heat load, so the refrigerant temperature rises. The TXV responds by opening and allowing an increased flow of refrigerant to the evaporator. A constant-speed compressor cannot accommodate the existing pressure lift and a new higher mass flow rate (because the TXV has opened), so the evaporator pressure rises (i.e., decreasing the compressor lift) and the increased flow rate is accommodated at the lower compressor lift. The new operating state point is at a higher evaporator pressure causing an associated higher evaporator temperature. Similarly, a decrease in cooling requirements leads to excess heat pump capacity. Of course, control systems are used to minimize the cold plate temperature variations, but if no thermal storage is used, the vapor compression heat pump must be sized for the worst-case, highest-capacity requirement.
Currently no thermal storage system can provide a significant improvement over the use of a larger fluid loop and radiator systems or vapor compression heat-pump-augmented spacecraft radiators. The difficulty with heat pump and fluid loop systems which have no thermal storage capabilities is encountered when they are used in thermal management applications having large differences between base and peak thermal loads. In the case of the heat-pump-augmented systems, the TXV senses an increased evaporator exit superheat when increased cooling is required. This occurs because the refrigerant flow rate is insufficient for the heat load so the refrigerant temperature rises. The TXV responds, as previously noted, by opening and allowing an increased flow of refrigerant to the evaporator.
Typical thermal transport system power dissipation requirements and duty cycles are base loads of 5 to 15 kW with 50 to 150 kW peak loads, and a duty cycle of 20 minutes at the peak load and 148 minutes at the base load.
It has been proposed previously to use a flexible expandable radiator as part of a vapor compression heat pump system. This approach makes, however, the heat pump reliability depend on a leak-free expandable radiator which has a number of potential mechanical drawbacks.
It is, therefore, an object of the present invention to solve the thermal storage problems by providing for efficient integration of refrigerant vapor storage within a vapor compression heat pump system.
It is yet another object of the present invention to provide a simple, compact method of storing heat pump cooling fluid.
In accordance with the device and method of the present invention, the heat pump adsorption storage device (HPASD), when full, would be discharged by using the hot exhaust compressor gases to raise the bed temperature and drive the refrigerant from the bed. This refrigerant would then be compressed, condensed, and stored in the liquid receiver for re-use. Availability of this approach is related to the ability of the storage device to rapidly adsorb or desorb the refrigerant.
Methanol makes a suitable working fluid for the 300.degree. K. evaporator to 375.degree. K. condenser application. When methanol is used as the working fluid, several adsorption materials are available. Synthetic alumino-silicates are made with a mixture of alkali methal hydroxides (e.g., Na or K) under pressure and high temperatures. The result is a crystalline structure with internal micropores capable of adsorbing polar molecules. Zeolites can adsorb 18.6 wt percent methanol at 325.degree. K. and if the bed is heated to 425.degree. K., the adsorption capacity is reduced to 2.6 wt percent. For this storage capacity, the vaporization of 1 kg of methanol at 300.degree. K. would store 1773 kJ of energy, but for a 16 net wt percent methanol storage capacity, 6.25 kg of zeolite would be required. The energy storage capacity in this instance is approximately 245 kJ/kg (1773/7.25=245). The actual storage capacity for this system is 52 kJ/kg when all factors are considered.